Hydraulic pump or motor



Jan. 29, 1957 E. c. DUDLEY 2,779,295

HYDRAULIC PUMP OR MOTOR Filed July 19, 1950 5 Sheets-Sheet l 9 ""0 i-- i Q 42 y (Q7 6% I/i'VENTOR.

g fi/i Jan. 29, 1957 ,E. c. DUDLEY 2,779,296

HYDRAULIC PUMP OR MOTOR Filed July 19, 1950 5 Sheets-Sheet 2 IN V EN TOR.

ATTORNEYS.

Jan. 29, 1957 E. c. DUDLEY 2,779,295

HYDRAULIC PUMP OR MOTOR Filed July 19, 1950 5 Sheets-Sheet 3 Jan. 29, 1957 E. c. DUDLEY 2,779,295

HYDRAULIC PUMP OR MOTOR Filed July 19, 1950 5 Sheets-Sheet 4 3 INVENTOR.

3 By? flow (415% ATTORNEYS.

Jan. 29, 1957 E. c. DUDLEY HYDRAULIC PUMP OR MOTOR 5 Sheets-Sheet 5 Filed July 19. 1950 INVENTOR.

/ ATTO/PA/E'rs.

United States Patent nrntgAuLic PUMP on oron Application July 19, 1950, fierial IS Io. 8 Claims. (Cl. 103-162) The present invention relates generally to hydraulic pumps and motors; and more especially to a hydraulic pump or motor operating at rela'tivelyhig'h pressures'in which two cylinderblocks'angularly disposed with re spect'to each other, rotate in a housing about intersecting axes and a plurality of angularpis'tohs interconnectth'e blocks and reciprocatein cyli ndrical'lb ores' inlthe cylinder blocks. 'A pump of this .t'ype'has comparativelyhigh efficiency. It may be used as a motor if fluid underpres sure is supplied to the unitjandrnay be used as a cornpressor if a'gaseous fluidis'useld. l

Pumps of this typea're comparatively old in the field of pump structures and have been proposed for many different. uses, but have actually never been widely used. This is probably a resultofihe fact that known devices arendt suited to delivery pressures of comparatively high magnitude, as forexatnple of the general orderof 500 pounds per sj qu are inch or higher. ,Rurnps of this geherallcliaraster have a metalto-r netal contact between the pistons and the cylinder walls/addicts impractical to provide packing or cylinder ringsfon the pistons, largely because of their small diameter-Q Known pumps of this type work very satisfactorily at low pressures of the order of pounds per square inch or less. .But when the delivery pressure exceeds these relatively low values by any substantial amounh'variousdiiiiculties are encountered, unless the designsarc changed.

in the first place, efficiency at ihiglrpressures is apt to be low because of a substantial amount of leakage of the fluid past the pistons and also around the cylinder-blocks. In pumps Where the housingis closed, a seco ndary result, anda far more serious one, occurs as a result of this leakage, which isthat hydraulic pressure is created that causes the cylinder blocks to lose their seal at theiren ds With The housing. Under this condition pressure or delivery rate iallsotlrapidly. Another difficulty with pumpsof known designs when operated at high pressurestis the fact that the metal bearing surfaces Wear with extreme rapidity because the high pressures are eccentrically applied to the cylinder blocks andnoprovisionhas been made to neutralize or balance these pressuresor to compensate for displacement of, parts under the'eccentric pressures.

Hence it becomesja general object of my invention to design a pump of thistype which may be used to deliver liquids atrclativelyihigh pressures forexample in the neighborhood otSDOupoundsper square inc hor higher.

It is also an object ofrnyiinvention to designahydraulic pump of this type in which the internal hydraulic pressures are balanced in such a manner as to prevent unduewear and maintain a high volumetric efficiency over along life of the pump.

It is a further objector my invention to design a pump of this character which is ablejto pump liquid'throughthe p p in t e td e io wit a minimumt tera i pf th pump structure and which may be also used either as a Pu pe amqtqtu Et r s ameles har bee status i ssumesr 2,779,296 fitt st! an- 2?,

structed according to my invention. in a pump of this character, a pair of angularly disposed cylinder blocks are rotatably mountedin a fluid tight housing] *The bedsing is closed at each end by an end plateover which'a fluid film is form'ed,and against which film the end surface of the cylinder"block-bears?'Both cylinder blocksare free to movsaxiany a limited amount within the housing under the infiue'nce of manure pressures applied to the blocks. Each block has a plurality of bores parallel to its axis of revolutionahd apluralityofangular piston members are slidably mounted in the bores of the we cylinder'bl'ock sf on 'e'nd" of piston member being in each cylinder block. The inner opposed-end faces of the "two" cylinder blocks "are imperfofate except for the cylinder bores, and are subjected to -fluid pressurefrom the working" fluid whichlfill s the spacebetvveen the inner faces of the blocks and within the housing. Thepunip is preferably designed to maintain the fiuidpressure, when operating normally, between the blocks at approximately midway between the "inlet and "outlet" pressures.

The net effective areas on the outer ends of the two blocks bearing against the 'fluid films over the adjoining end plates are equal for both blocks and are designed to have a particular riet area with respect to the net 'areas of the inner ends. As a consequence, the forcedue to fluid pressure at the inner endsoffthe blocks is able to firmly seat the cylinder blocl rsagainst the fluid film over the end plates maintain afseal between the cylinder blocks and the housiug overfthe normal range of operating pressures. -If the bea'ri g area atthe outer end of the cylinder block is too large m proportion, then the force due re fluld pressure atitheinner'end of the bl'oolcis insufii 'cient to seat "the 'bloclcafnd the fi seal between the block and end plate is brokenwithfth resultthat the desired delivery pressure canno t be maintained Converse ly,' it this outer endarea istoo srnal hebea'ring pressureagainst the end plate becomes so and the cdntactin "metal surfaces wear with extreme rapidity. Ezrperienoe-showsithat the net area on the inner end of eaohblock subjectedto hydraulic pressure should be about 11. 6 5 to1:95 tinies the area cnthe outer end of the block incontaq't wfthitheifluid film over the end plate.

a preferredjmeans j fon controlling the total tor'ce applied uh drauli'callyto'the outer end of the piston, each piston blockhas its outer end centrally recessed, thetwo recesses having aneduahdiarn'eter at the outer face of thecylindejrblocks'. Bothflendplates are then provided withsuitable fluid passage means extending ,from said recess either directlyprindirectly to the'fluidinlet passage so that hydraulic pressure within the area ofthe recess does not exceed the hy'draulic pressure at the inlet of the pump. By controlling the diameter of' the se recesses and leaving other dimensions constant the total force on the oute'r end of eachblock canbe controlled whicliinturn afiord's a, means for closely balancing the force dueto fluid pressure against theinner ends of the cylinder bloclgs. i i i How the above objects and advantages of my inven tioh, as well asjoth ers not' specifically-{set forth, are" attained wil be" more jreadilyunderstood by reference to theffollo g d'escription'and to the annexed' drawings in Fig.1 is a plan viewof a pump constructed accordin Fig. 4 is anje'ndl elevation n Q L Fi -Z; c

. t al' 'ruediansection on line S- S of the pump viewed from arran e Fig. 6 is a vertical median section on line 6-6 of Fig. 4;

Fig. 7 is a transverse section on line 7-7 of Fig. 6;

Fig. 8 is a view on line 8-8 of Fig. 6 showing one end plate in elevation;

- Fig. 9 is a view on line 99 of Fig. 6 showing the other end plate in elevation;

Fig. 10 is an enlarged view on line 1010 of Fig. 6 showing an end of one cylinder block in elevation;

Fig. 11 is a perspective view of a single cylinder block removed from the housing with three angular pistons in the block, the remaining pistons being'removed;

Fig. 12 is a perspective view of a single hollow piston member;

Fig. 13 is a horizontal median section similar to Fig. of the variational form of pump;

Fig. 14 is an elevation of one end plate as indicated by line 14-14 of Fig. 13;

Fig. 15 is an elevation of the other end plate as indicated by line 1515 of Fig. 13;

Fig. 16 is an enlarged fragmentary section on line 16-16 of Fig. 15;

Fig. 17 is a side elevation of another variational form of fluid pump;

Fig. 18 is a horizontal median section on line 18-18 of Fig. 17; and

Fig. 19 is a transverse section through one end plate on line 1919 of Fig. 17.

As shown in Figs. 1 to 6, the pump has a fixed body which consists primarily of a central housing which is indicated generally at 10, and end plates 11 and 12 at either end of the housing. Housing is preferably made as an integral unit, but it can alternatively be made in sections bolted together. The housing has a central cavity formed in two sections, each section consisting of a cylindrical bore 14. The two cylindrical sections 14 are angularly disposed within housing 10 in such a manner that their axes intersect, as shown in Figs. 5 and 6. The interior angle between the two axes is here shown as being 150"; but it will be understood that my invention is not necessarily limited to any particular angle since a greater or lesser angle may be found better adapted to a particular situation. Of course this angle cannot be less than 90 and the pumping action of the device is reduced to zero when the angle between the axes reaches 180, at which point the axes become coincident.

Within each cylindrical section 14 of the housing cavity is a cylindrical block which is rotatable inside the housing, the axis of the cylinder block coinciding with the axis of the cavity section 14. One such cylinder block 15 is shown removed from the pump housing in Figs. 10 and 11.

As may be seen from Figs. 5 and 6, bores 14 terminate at their inner ends in a short section of slightly reduced diameter, creating an annular shoulder 17 at the inner end of each of the two angularly disposed bores 14. The two shoulders 17 each provide a means for limiting the inward movement of a cylinder block 15 within the housing. Because of the spacing between the two shoulders 17, these shoulders also provide means for separating the adjacent or inner ends of the two cylinder blocks in order to expose the inner ends of the blocks fully to the pressure of working fluid, as will be described more fully. At their outer ends, each cylinder block 15 bears against a fluid film over an end plate 11 or 12 when the pump is in operation. In my construction, each cylinder block is free to move a limited distance axially toward and away from the end plate. The amount of this axial movement may be very small, perhaps only a few hundredths or even thousandths of an inch, and is therefore not indicated in the drawings. However, it is sufficient that each cylinder block may be characterized as being unrestrained axially in a mechanical way for such range of movement, thus permitting the axial position of the cylinder block under operating conditions to be determined by the forces applied to the cylinder block by fluid pressure. As will be again discussed below, the axial movement of each cylinder block allows a film of working fluid to be established and maintained between the block and its adjoining end plate. This film not only provides lubrication at the outer end of the block but exerts a pressure inwardly on the outer end of the block. The bearing areas on the outer end of the block actually bear directly against the film and are not in metal-tometal contact with the end plate when proper operating conditions are maintained. In this condition the block is referred to herein as being seated against the end plate since the block is forced outward toward, but is properly spaced from, the end plate which ultimately resists the outward thrust of the block.

As may be seen particularly in Figs. 10 and 11, each cylinder block 15 is provided with a plurality of cylin drical bores 18 which are all parallel to the axis of the block and to each other. In each of these bores is an angular piston member 20, the angle between the two legs of the piston member being equal to the angle between the two sections 14 of the housing cavity. One end of each piston member 20 fits into a bore 18 in one block 15, while the other end of each piston fits into a corresponding bore in the other cylinder block. The piston members drivingly interconnect the two cylinder blocks so that as one block is rotated by power applied thereto, the other block is rotated at the same speed by force transmitted through the pistons. As the blocks rotate, pistons 2t) slide in and out of bores 18.

In the preferred form of my invention illustrated in Figs. 1 to 12, piston members 20 are hollow sleeves. However, my invention is not necessarily limited thereto and pistons in the form of solid rods are illustrated in a modified form of my invention described later. In either case, the operation of the pumps is essentially the same. When piston members 20 are hollow sleeves as shown in Fig. 6, the pistons serve to connect the two bores 18 in the two blocks to form in effect a continuous space the volume of which varies from a minimum which is substantially equal to or slightly greater than the internal volume of piston 20, as shown by the top piston of Fig. 6, to a maximum volume which is enlarged from the minimum by nearly twice the volume of a bore 18 between the end plate and the end of a fully retracted piston 20, as shown by the bottom piston of Fig. 6. Since there are nine piston members 20, there are nine pumping chambers which change in volume to effect the pumping action. When the pistons are solid rods, the variable volume is that space between the end of the piston and the end plate so that there are then eighteen such chambers the volume of which is determined at all times by the spacing between a fixed end plate and the movable end face of a piston. As will be apparent from the draw ings, the minimum volume of any one such chamber may approach zero and at most is but a small fraction of the maximum volume. K

Under many circumstances I have found it advantageous to employ hollow pistons, and especially so because the pressure of the working fluid inside the pistons tends to expand them against the walls of bores 18, thus reducing the tendency for working fluid to by-pass the plstons in the clearance between the pistons and the bore walls. In addition to improving the volumetric efliciency of the pump, this condition eliminates the critical nature of extreme precision in forming the cylinder bores and machining the pistons. There is also a very substantial reduction in the total thrust against the ends of the piston, a thrust which tends to buckle the pistons at their angular mid-point and which tends to make the pistons bind in the bores. Reduction in end thrust reduces friction and power consumption and improves the life of the P p- It will be noted that the two end plates 11 and 12, which are attached to central body housing 10 by bolts 21 or any other suitable means, are not duplicates. End plate 11 is provided with two fluid passages, inlet passage 23 and outlet passage 24. Adjacent an outside surface of end plate 11 inlet passage 23 is threaded as at 25, or otherwise formed in a suitable manner, for attachment to pipe 26 delivering fluid to the pump. In a similar manner outlet passage 24 is threaded or otherwise finished at 2'7 for attachment thereto of conduit 26a which is the delivery line by which fluid under high pressure is discharged from the pump.

Inlet passage 23 and outlet passage 2 both open to the bearing or inner face of end plate 11 against which the cylinder block 15 bears. At this face, inlet passage 23 terminates in arcuate inlet port 28 while outlet passage 24 terminates in a similarly shaped arcuate port 29, as shown in Fig. 8. The radial width of each port 28 and 29 is equal to the diameter of cylinder bores 13, the ports being shaped to conform to the path of the bores as blocks 15 rotate. Thus, bores 18 may receive the fluid from inlet port 23 during the period that they are at least partially open to the port and discharge fluid to port 29 for such time as they are at leastpartially open to that port. The adjacent ends of the two ports are separated by a distance slightly greater than the diameter of a bore 18, and preferably equal to the diameter of one bore plus the minimum thickness of two Webs between bores, in order that the bores are completely sealed momentarily by the end plates between the suction and compression strokes of piston 2t} and cannot communicate with both the inlet and outlet ports at the same time.

Inlet port 23 is preferably slightly longer than outlet port 29 in order to make for smoother operation and better filling of the cylinders on the suction stroke. This can be done since flow is always in the same direction through each port. If the pump is made reversible in a way to reverse flow through a given port,'then it is better toinake both ports of the same length and symmetrical about a median plane vertical in Fig. 8 so that operation is then the same for either direction of revolution of the cylinder blocks.

End plate 12 is provided Withfluid circulating passage 3%) which terminates at one end at inlet port 31 in that face of the end plate against whicha cylinder block 15 bears. Port 31 is similar in size and. shape to inlet port 28 and occupies the same angular position relativcto the axis of rotation of the adjacent cylinder block. The other end of fluid passage 3% communicates with a longitudinally extending fluid passage 33 in central housing iii, which fluid passage at its other end communicates with branch 23a of fluid passage 23 in plate 11. By this arrangement, fluid entering the pump can flow through passages 23a, 33 and St to enter the pumping chambers through both ports 2% and 31.

in order to balance the forces exerted by fluid pressures on the two cylinder blocks, as will be more fully explained, it is desirable to provide end plate 12 with recess 32 which has the same size and shape as outlet port 29 and is, in eifect, a dummy outlet port in end plate 12 located at the same angular position with respect to the adjacent cylinder block as is outlet port 29 relative to its adjacent cyiinder block. The size and location of recess 32 must be as carefully determined as any of the other ports, except that its depth is not critical .provided it is suificiently deep that there is no fluid film pressure on he cylinder block over the area of this dummy port.

The advantage of using two inlet ports and a single outlet port is that this arrangement insures circulation of fluid through the pistons and eliminates the possibility that a small quantity of fluid can be trapped for a considarble period of time in the interior of the pistons. stagnant i'luid rapidly becomes. heated, as a resultxof the pumping action, and reaches a relatively-high ;tem perature if it remains in :the pump for any appreciable length of time. By having: cool liquid. at; the. inletnside enter into both ends of each compression chamber, all the liquid is scavenged and forced out at the outlet side before it becomes overheated. it will be apparent that the same result can be achieved by other arrangements, as for example, using two outlet ports and a single inlet port. However, the arrangement shown is preferred since the greater area of fluid passages is provided on the low pressure side of the pump, thus assisting in maintaining the pumping chambers completely filled at all times, which is important with the pump rotating at high speeds.

In some pumps .of this general type, it is possible to reverse flow through the pump merely by reversing the direction of rotation of the cylinder blocks. However, that is not possible in this construction since the various fluid passages and ports are not all symmetrically located. However, it is possible to reyerse the flow through this pump very simply by rotating both end plates 11 and 12 with respect to housing 1 and by providing in body 10 a second fluid passage -35 which connects inlet passages 23a and 30 in the end plates after they have been rotated .180". it will be noted that after this change has been made fluid still enters the pump through passage 23 and is discharged through delivery passage 24; but these passages have been reversed in position from that shown in Fig. 5 so that fluid flows from conduit 28 through the pump to delivery pipe 26. The provision of passage 35 duplicating passage 33 in the pump housing permits this change in direction of flow with a very simple change in the pump assembly and without altering the position of housing It) which'is normally fixed in position since it is bolted or otherwise attached to some base, piece of machinery,.or the like.

The pump is driven by drive shaft 37 which is connected at its outer end to an electric motor or other suitable type of prime mover, not shown in the drawings. One end of drive shaft 37 extends into a central bore or recess 33 in one of the cylinder blocks. A driving connection is effected-between the shaft andthe cylinder block by means of a ,key 39, or other suitable type of connection. A key or spline connection is preferred in order that shaft 37 can be held against longitudinal movement and still not restrain the driven cylinder block against the limited axial movement described above since this type of connection permits relative paxial movement between the driven cylinder block and the drive shaft. Alternatively, the cylinder block and drive shaft can be rigidly connected together but the drive shaft permitted to move a limited distance axially in order to seat the cylinder block properly under operatingfluid pressures.

Suitable bearings for shaft .37 are indicated at .40, the bearings rnounted in end plate '12. Bearings 40 are held in place in the. end plate by hearing cap141 which is provided with an oil seal 42, ofany conventional type. The particular tpump illustrated herein is designed to pump a lubricating liquid .which may be used to lubricate bearings :40, astwill be-explained,andsealsil prevents leakage of such lubricant out Ofthe pump. a In the event that \a non-lubricating liquidis being pumped, other arrangements may be made, tolubricate bearings 40, in which case a vdiferent type of oil seal may be. required.

Whema lubricating liquid isbeing pumpedby the unit, it is convenient to allowliquidfrom the inletfluid passages to. reach bearings fitl in order-tolubricate them. This is done conveniently by opening fluid passage-30 directly to the chamber containing bearings 40, as shown in Fig. 5. This lubricating-.opening is,in the opposite side of the end plate fromportfil, andissmallerin size. In this way the bearing chamber is filled with liquid at the same pressureas:ismaintained in'the inlet passages of the pump; and thissameypressure relationship iszalways maintained because. of the-freecommunicatipn with the inlet-passage at. 30.

,vAs shown; pa rtic larly inglfiigs j and d the other ylinder bloc Ihf u s; th n rwhit ri$z 9tz fitmn etg directly to drive shaft 37, is also centrally recessed at its outer end at 44. The diameter of the two recesses 38 and 44 are made equal at the outer end faces of'the two cylinder blocks. Since shaft 37 may have a smaller diameter than the desired diameter of recess 44 when determined by fluid pressure considerations, recess 38 may be counter-bored at 3801, the counterbore being the same diameter as recess 44. This maximum diameter of the two cylinder block recesses measured at ends of the cylinder blocks is preferably approximately 1.9 times the diameter of cylinders 18. A variation of approximately ten percent either way from this value is permissible under some circumstances although it will be appreciated that the diameter of the recess cannot be increased greatly.

In order to prevent fluid pressure from building up within recess 44, fluid passage 43 is provided which extends from recess 44 to fluid inlet passage 23. Conse quently fluid pressure in recess 44 is the same as pressure of the fluid in the inlet passage on the end plate adjoining recess 44. In order to similarly limit the fluid pressure in recess 38, end plate 12 is provided with passage 45 extending from counterbore 38:; into the space occupied by bearings 40. As explained above, fluid pressure in the bearing chamber is at the inlet pressure and any excess above this pressure in the cylinder block re cess 38 is relieved through passage 45. Passage 45 is here shown in the form of a groove in the wall of the bore in the. end plate through which drive shaft 37 extends, the groove extending axially of the shaft. it will be evident that in lieu of the groove, the bore through which shaft 37 passes may be enlarged sumciently to accomplish the same result of relieving excess pressure at the end of the cylinder block. However, the diameter of the bore in the end plate must not exceed the diameter of counterbore 38a.

If a non-lubricating liquid is being pumped and it is desired to exclude this liquid from the chamber con taining shaft bearings 40, pressure relief can be obtained with a simple change. Inlet passage 30 is closed to the bearing chamber by making the outer wall of end plate 12, opposite port 31, imperforate. Relief passage 45 is likewise closed off to the bearing chamber. ln order to obtain an adequate seal, conventional types of oil seals may be employed in addition to those illustrated. iassage 45 is then relocated to extend between counterbore 38a and inlet passage 30 so that excess pressure in the central recess of the cylinder block is relieved directly into the inlet passage in the same manner as at the. opposite end of the pump by passage 43.

As a part of the system for controlling fluid pressures within the pump, housing 10 is provided with a longitudinally extending fluid passage 46, shown in Fig. 6. Passage 46 is in the form of a groove in the wall of the two cylindrical bores 14 forming the internal cavity in the housing. Passage 46 extends from end-to-end of the housing, and opens in the middle to the central space between the inner opposed ends of the two cylinder blocks. Fluid passage 46 is in the nature of a pressure balancing passage since its purpose is to regulate or control fluid pressures in the housing rather than to permit flow of any appreciable quantity of fluid. This passage is located in housing 10 to come at a point mid-way between the ends of inlet and outlet ports 28 and 29 respectively.

During the suction stroke while pistons 20 are being retracted into cylinders 18, fluid flows into the cylinders from ports 28 and 31. During the compression stroke when the pistons are advancing toward the ends of the cylinders, liquid is forced out of the cylinders and into delivery port 29. In the ordinary course of operation, fluid leaks around the pistons and around the cylinder blocks into the central space within the housing cavity between the inner ends of the two cylinder blocks. This leakage occurs normally since the fits between the moving parts are not sufliciently-close to prevent it. The parts may be fitted according to the current standards of mafit chine shop practice which provide for a diflerence in diameters of bearing members of the order of one-thousandth of an inch per inch of diameter, with a normal minimum of substantially one-thousandth inch.

However, in order to make sure that this inner space fills quickly and at the desired pressure, fluid passage 46 is provide which permits fluid to flow around the cylinder blocks with comparative case from the outer ends into the space between the two blocks. As a consequence, after the pump has been running a short time, this central space is filled with working fluid under a pressure which is preferably maintained at a value about mid-way between the inlet and outlet pressures in passages 23 and 24 respectively. Although there is a natural tendency for the pressure in the fluid at this location to reach an equilibrium value in the vicinity of the mid-point between inlet and outlet pressures, because the leakage path into this central space from a cylinder under high discharge pressure is approximately equal to the leakage path into a cylinder under low inlet pressure, nevertheless the value of the fluid pressure between the cylinder blocks can be controlled within limits.

Since the inner ends of the two cylinder blocks are imperforate, except for the cylinder bores which are sealed off by. the pistons, fluid pressure is exerted against these blocks over substantially their entire inner end faces. The net area subjected to fluid pressure on these faces equals the area of a circle whose diameter is the outside diameter of the piston less the total cross-sectional area of the pistons. This pressure of the Working fluid exerts a force axially of each cylinder block which presses it againgst the bearing seat formed by the inner face of the associated end plate 11 or 12. The reaction to this thrust is provided by an opposing force developed by the fluid film between the end plate and the cylinder block or, in the event the fluid film fails or in discontinuous, the pressure of the end plate itself where there is metal-to-metal contact.

it will be apparent that the magnitude of this reacting force supplied by the fluid film on the outer end of each cylinder block is a function of the fluid pressure and of the area of the cylinder block bearing against the end plate. First, consider the areas involved. Since the net area on the inner end of each cylinder block subjected to a common fluid pressure is equal to the area on the other block, it is at once apparent that the net effective area on the outer end of each block bearing against an end plate should be equal to the similar area on the other block bearing against the other end plate.

In addition to the equality between corresponding areas on the two cylinder blocks, it has been determined by experience that the net areas on the inner and outer ends of the same cylinder block should also have a definite relationship to each other. This is a prerequisite to establishing a resultant force on each cylinder block which firmly seats the block against its end plate but which is not excessive for this purpose. The result may be expressed as a balancing, but not an equalizing, of the axially applied hydraulic forces upon each cylinder block in order to keep the block seated as it rotates and thus maintain a fluid seal at this end of the block. Experience has indicated the following general rule for the relationship between the two end areas of the cylinder block: The net area at the inner end of the block subjected to fluid pressure is ideally about 1.75 times the net effective bearing area between a cylinder block and its end plate. Actually, some deviation from this ideal value is permitted and satisfactory result may be obtained with a value for this factor from 1.65 to 1.95. v

This relationship of areas can be differently stated in the case of the preferred form of pump illustrated in which the radial width of the inlet and outlet ports is equal to the diameter of cylinder bores 18 and in which the distance between the adjacent ends of these ports is not greater than the diameter of one bore plus two times thejmihirhtim thickness state weeps It is not necessary that these dimensions be used; but when thesqpreferred dimensioiis exist, the netarea at the inner'end of the block subjectedto'fiuid pressure is ideally two jtimes the netetiective bearing are'aas "determined in the following paragraph Again, satisfactory results may beo'btained with some deviationsfrdin'this value but such deviation should not exceed approximately percent of this factor. Y H y Y I In determining the net eifec'tive hearing area on the outer end are cylinder block it is db'vious that there can be subtracted from the gross dross se'dtidhal area of a block (e t1ual to the area of a circlehaving adiameter equal to the outside fdiameterof mee' uader block), the cross sectional meet thecentr'al recess Lat '38:} or 44. It is also foundias a mattero'f experience that there can be subtracted the total area of what is "herein termed the piston 'ahnulu s. This is the admins on the end of the cylinder block between circles 15a and 15b in Fig. 10 in Which all the cylinderbore's 1 8 lie aiidyvhichhaS a r-adial width equal to the diameter pr pores is.

Although there is an appreciable afeaon the cylinder block lying 'within the inner andouter circles defining this annulus and between the 's'iiecessive cylinder bores, it will "be understood that only "asinall part of this area is at any given inst-ant facing an end plate. This is true because ports 28 and 29 in one end plate and port 31 and recess 32 in the other end platecoincide with the piston annulus over much of the latter. jHence, it has been found as a matter of actual practice that the various areas of the pump can be designedfon the assumption that one can ignore any bearing between the block and end plate within the area of the piston aiiriillus.

Consequently, the net 'eifective area on the block in bearing at its outer end is practieally limited to two annular areas. One is the annulararealying outside the piston annulus but inside the outer "diameter of the cylinder block; the other area isa smaller one inside the piston annulus and outside of thebentral recess 38a or '44. For a given cylinder block the total area in bearing on thefend of the block canibe altere d most easily by changing the diameter or the centraljecess If the diameter of the recess is decreased, the total amount of bearing area increases and, assuming a coiietaht fluid pressure in the filth at the outer end o'f th e bloclr, a consequently increased force is javailable at tlie end of the block to resist the fluid pressure thrils't at the 'inner :end of the block. This condition allows a "larger fluid pressure between the two cylinder blocks to build up; while the fluid pressure between the cylinder blocks is decreased by increasing the diameter of the center recess to decrease the bearing'area and the total reaction on the end of the cylinder block. V I e d d For a given bearing area on the outerend of the cyl-. inder blocks, the maximum fluid'pressurebetween the two blocks 'is controlled by the pressure in the fluid film maintained between the cylinder blocks and the end plates. The fluid pressure between the blocks seats them against the end plates and so effects "a seat between the cylinders and the end plates. If mantra pressure reaches a sufiieiently high value, the inward axial thrust on the cylinder blocks is greater than the outward, thrust and the blocks are unseated, with a loss of the liquid tight seal. The resultant loss of fluid from the pumping chamber's causes a loss of delivery pressure and a very marked drop in pumping efii ciency. On the other hand, if the pressure in the film for any reason drops too low, then the outward thrust on the blocks from the center pressure is able to break down the liquid film and metal-to-rnetal contact results with a-high rateof we'ar and poor-pumping efiiciency. Consequently, the'ideal 'situationi's one in which the liquid film pressure is maintained at the proper value to exert a thrustwhich exactly'balances the ontitvard'thrust'applidto the-'iniier'endsof the'twoblocks.

Obviously this total thrust ea be than may whitened by controlling the teen bearing area an the ends of the cylinders, and it will now be explained how these areas are designed. l

It has been found that a preferred diameter (1. D. in Fig. 10) for recess 44 or counterbore 38a is equal to 1.3 times R (as defined below) times the diameter of cylinder bores 18, indicated at in Fig. 10. This value of I. D.=1.3RB can be varied somewhat since it is not a particularly sensitive va lue, and can be changed as much as ten percent either way without necessitating a s ignficant change in other dimensions. However, the value of 1.3RB for thediameter of the centralrecesses is preferred since it maintains the working pressure between the two cylinder blocks at substantially the average pressure between theinlet andoutlet pressures. This central pressure can be increased ordecreasd by as much as ten percentby controlling the bearing area on the end of each cylinder block, but keeping the areas on the two blocks equal at all times. However, any substantial departure from the preferred value results in a decrease of volunietric efficiency, as well as other diifictilt ies.

In designing a pump, the first step is to establish the rate of dischargein gallons per minute. Since the de livery rate is a function of the speed of revolution of the cylinder blocks, this variable quantity is eliminated by basing all computations on a delivery rate at some standard speed, as fo r example 1200 R. P. M. On this basis is then possible to calculate the total displacement per revolution required of all cylinders 18. This displacement per revolution is the same regardless of whether piston'members 20 are hollow or solid, provided that the discharge from both ends of the blocks is combined in the ease of solid pistons, as will be pointed out more fully later. l l

in order to determine the displacement of each cylinder, the total displacement per revolution is divided by the number of cylinders. This number is preferably an odd number greater than 8. It has been found that 9, 11 or 13 cylinders provide a substantially continuous delivery from the pump without perceptible pulsations. In the drawings the cylinder blocks are shown as having nine cylinders each, which is a'preferred number. Having selected the number of cylinders, the cylinders are laid out with a size or -diameter B and at a radial distance from the block axis such that two conditions are satisfied. First, the stroke of the piston times its cross sectional area equals the desired displacement per cylinder. Second, the area of the piston annulus, as defined above, bears a desired ratio to the "total cross sectional area of all the cylinders. To meet the latter condition, the cylinders are placed in the block with their centers on a circle'lsc having a diameter C, termed the piston circle, thedimensions being such that the total area of the annulusis greater than about 1.2 times the total cross sectional areaof all cylinders 33, but less than about 1.7 times thecross sectionalarea of all cylinders. Expressed mathematically, the ratio (R) of the annulus area(A) to the total'area of the cylinders (11a, when 12:110. of cylinders and. a= -area of one cylinder) lies between about l.2'and 1.7. If the value of this ratiofalls beiow 1.2, walls between successive cylinders become too thin to withstand high pressure, while if the value of the ratio greatly exceeds 1.7 the area of the cylinder block between successive cylinders becomes so greatthat a film pressure can be established for at least a part of a revolution between each pair of cylinders. Under this condition,'the basic assumption that'there is no film pressure over thepiston annulus is -no longer valid.

Having established the piston annulus, the diameter C ofthe piston-circle 15c-indicatd in Fig. 10, is established. Also, the diameter-'of'the 'ceiitral recesses is established, since the values R ai1'd have-been' established. With these vanes resumeir aaw'passiba to d'eterinine the r rasse 11 diameter of the outside of the cylinder block which is calculated according to the following formula:

All the variables in this formula are identified as follows:

A=gross area of the piston annulus a=area of a single cylinder bore n=number of cylinder bores B=diameter of each cylinder bore k=arbitrary constant having a value of from 4.2 to 4.4

inclusive The above formula and other area relationships are purely empirical and have been found to produce a pump for liquids which operates over a wide range of pressures. Pumps designed according to these standards ranging in size from one gallon per minute to forty gallons per minute, have operated at speeds up to 3600 R. P. M. and at pressures up to 4000 pounds per square inch. By experience it has been found that the features of successful design for a given pump can be translated into terms of pumps of other sizes or capacities by the above formula. The constant It has a preferred value of 4.34 but may vary from 4.2 to 4.4.

The above equation has the effect of increasing the radial dimension of the outer annular bearing area at a rate somewhat greater than the rate of increase in the diameter of the cylinder block. If the value of k is too great, the radial dimension of this annular bearing area becomes too large and the total bearing area of the piston block is so great that the film pressure prevents the block from seating properly against the end plate. When the block does not properly seat, undue leakage past the cylinder block is permitted, and volumetric efliciency drops as well as causing undue wear as the result of irnproper positioning of the cylinder block. The reverse situation occurs if the value of k is too small because then the end bearing area on the cylinder block is too small to provide adequate support for the block and excessive Wear takes place.

Since the ratio R is a ratio of areas, the above equation can be rewritten in terms of R. Substituting the preferred value for constant k the above formula becomes: O. D.=B(Rl+4.34) which expresses the diameter of the cylinder blocks in simplified and preferred terms.

A pump of preferred construction designed according to the above standards has cylinder blocks with the desired area ratio. It will be found, possibly after some trial and error calculations, that the area on the inner end L of each cylinder block exposed to the central fluid pressure is approximately 1.75 times the bearing area on the outer end of the cylinder block. For best results the ratio of these two areas should be maintained within a range from about 1.65 to about 1.95. The ratio of these areas maintains a balance between the two axial thrusts upon the cylinder block mentioned above to keep the block seated at all times.

When in operation, working fluid under pressure reaches the central space in the housing between the two cylinder blocks, groove 46 being provided to facilita'te the flow or fluid into this space. Normal clearance around a block 15 is sufficient to permit fluid to reach this space and eventually to maintain it at the desired pressure; but the provision of some specific fluid passage such as 46 is preferred in order to fill the space quickly and surely and maintain the desired pressure on the inner ends of the cylinder blocks. These cylinder blocks are spaced apart at their inner ends by virture of engagement with annular shoulders 17, thus presenting the entire ends of the blocks, less the areas occupied by piston members 20, to fluid pressure. This fluid pressure causes the blocks to be seated firmly against the end plates at the ends of housing 10; and in order that the blocks may seat properly under fluid pressure they are held substantially unrestrained in an axial direction in the housing, although they may move only a very limited amount. The force exerted by fluid pressure in the central space on the inner ends of the blocks is balanced by the force exerted by the fluid film pressure against the bearing area on the outer end of each block. This bearing area is closely controlled in order to provide an adequate bearing area so that the unit pressures are not high enough to disrupt or break the fluid film, but at the same time keep the total force inwardly on the block at a value low enough that the block is not unseated from the end plate. As a means for controlling the fluid forces against the outer ends of the blocks, the blocks are centrally recessed, these recesses having equal diameters at the outer end face of the block. Fluid passage means is provided at 43 and 45 to relieve the pressure in this recess and maintain it at a value equal to the inlet pres sure of fluid, which is the pressure maintained in inlet passage 23.

It is obvious that the fluid pressure on the inner ends of the cylinder blocks is uniform over the entire area of the block. The resultant force is concentric with the block. This is not the case at the outer end since the fluid pressure in the vicinity of outlet port 29 is higher than in the vicinity of inlet port 28. Consequently, the resultant of the reaction on the outer end of the cylinder block is eccentric with respect to its axis of revolution. The eccentric force causes the cylinder block to tilt slightly so that its axis of revolution is not precisely coincident with the axis of the housing. The amount of this inclination depends upon the normal clearance between the cylinder block and a wall of the housing cavity, but in a typical pump would be of the order of .001 to .002 inch per inch of length of the cylinder block. While this inclination is small it concentrates wear on the end plate in the vicinity of the inlet port. A refinement of construction which is optional may be adopted to compensate for this condition. The inclination of the cylinder block is calculated and then the bearing face of the end plate is tilted a corresponding amount with respect to a plane exactly normal to the axis of the adjoining section of the housing cavity. The result is to achieve parallelism between the end of the cylinder block under operational loads and the face of the end plate so that bearing between these two members is distributed more evenly over the entire bearing areas. While under some circumstances this refinement in manufacture might be of little consequence, it is significant in the present pump because of the relatively small bearing areas involved.

There is illustrated in Figs. l3, 14, 15 and 16 a modified form of a pump constructed according to my invention which is adapted to be driven in either direction in order to reverse the flow of fluid through the pump without the ecessity of changing or altering the location of any of the pump parts. It will be noted that inlet fluid passage 23 and branch passage 23a are the same as in the form of Fig. 5. While the outlet passage 24 is the same as previously described, there has been added to end plate llla branch outlet passage 24a, similar in shape and location to passage 23a. Passage 24a is always in communication with passage 35 in housing 10 which is as described above. In the form previously described, passage 35 duplicated fluid passage 33 but only one of the two passages was used at a time, the other one being available in the event it was desired to reverse the pump. In the present form, both of the passages 33 and 35 are used at all times, passage 33 being a part of the inlet passage system and passage 35 a part of the outlet passage system. Thus the outlet fluid passages are similar to and symmetrically disposed with respect to the inlet fluid passages in end plate 11a.

A similar change has been made in end plate 12a to which an outlet passage 3011 has been added which is symmetrically disposed with respect to inlet passage 30.

Passage 30acom1nunicates with passage 35 in'th'e pump housing and "also with outlet port 31a which takes the place of recess 32 of Fig. 5. Each'end plate 11:: and 12a is thus provided with an inlet port and an outlet port which are of equal length and symmetrically disposed with relation to a median plane.

The passage means for relieving pressure in central recess 44 has also been changed. It consists of passage 50 opening into recess 44 and having two branch passages 50a and 59b which communicate respectively with inlet a'n'd outlet passages 23 and 24. In each of these passages is a ball check valve 51 and 52 respectively. When the pump is operating in what may be termed the forwarder normal direction, fluid "enters through passage 23, which is at correspondingly low pressure. Fluid leaves through passage 24 which is-atcorrespondingly high pressure. The high pressure closes valve 52 and is prevented from entering passage 59b by the valve. Any excess of pressure in recess '44 over the inlet pressure opens valve 51 and the pressure is relieved through passages 50 and 58a in the mariner previously described. In the event that the direction of the pump is reversed so that fluid enters the pump through passage 24 and leaves through passage 23, it will be apparent that check valve 51 automatically operates to seal olf high pressure passage 23 from the cylinder block recess and yet excess pressure in the recess can be relieved through passage 50b and valve 52 into the low pressure inlet passage 24.

Corresponding changes have been made in the other endplate Ma. The direct connection between inlet passage 30 and the chamber occupied by bearings 40, as shown in Fig. 5, has been eliminated and replaced by passage 53 which contains check valve 54. A similar passage 55 containing a check valve 56 has been added placing outlet passage 30a also in communication with the bearing chamber. However, when the pump is "operated in the forward direction, the high pressure in outlet passage Twila closes check valve 56, preventing any loss of high pressure fluid into the bearing chamber. At the same time any excess pressure reaching the bearing chamber through relief passage 45 opens check valve 54 and allows pressure to be relieved through passage 53 in the same manner as in the form of Fig. 5. It will be clear that in the event the direction of flow through the pump is reversed, that high pressure closes ball check 54 and allows excess pressure to be relieved through passage 55 when the pressure is sufiicient to open check valve '56 which is now opposed only by the low pressure on the inlet side of the pump. p

The four check valves are here shown as being ball check valves which are typical of the kind of valves suited to this use. Of course any other suitable type of check valve can be used instead.

From the foregoing it will be seen that flow of fluid through the pump of Figs. 13-16 may be reversed merely by changing the direction of rotation of shaft 37 and without changing the position of any of the pump parts, and more'especially that it is no longer necessary to rotate the end plates through 180". The duplication of certain fluid passages which are now disposed symmetrically in the pump and the addition of dual ball check valves in these pressure relief passages has permitted this change. Except as noted, this form of pump is the same as the one already described in detail.

A further modification of my invention is illustrated in Figs. 17, 18 and 19. In this form of pump, one basic change is that Z'pisto'n members 201") are solid instead of hollow as in the forms already discussed. Making the piston members solid has the effect of dividing each pumping chamber into 'twe separate parts, but the total dis placement effected by each piston member for each stroke is the same as before.

With solid pistons 24)]; it is no longer possible to introduce and discharge fiuid froth one 'end of the en-sip only as before, but it is now necessary to introduce and deliver fluid from both ends of the pump since the fluid entering one cylinder block discharged from that block at the same end of the pump. Under these circu'm stances, it is convenient to provide-housing 1% with two longitudinally extending fluid passages 33b and 35b which serve respectively as inlet and outlet manifold passages and are comparable to similar passages previously described. Housing 1011 is provided with a threaded nipple 25b by which a fluid conduit may be connected to the housing in communication with inlet passage 33b; and a similar threaded nipple 27b is provided on the opposite side of the housing in order to permit connection thereto of a fluid conduit in communication with fiuid delivery passage 3%.

Each of end plates 11]) and 12b is provided with a fluid inlet passage 60 or 61 respectively. The two fluid inlet passages are duplicates of each other and terminate at ports 64 which are located at the same position angularly with respect to the axes of the cylinder blocks. Similarly, the two end plates are provided with duplicate outlet fiuid passages 62 and 63 opening at ports 65 which are also similarly located with respect to the axes of cylinder blocks 15. Each inlet passage 60 or 61 in the end plate communicates at one end with fluid passage 3312, while the other end of the passage terminates in an arcuate inlet port 64, one of the two similar ports 64 being shown in Fig. 19. In a similar manner, out let passages 62 and 63 each communicate at one end with fluid passage 35b in the central housing, and at their other end terminate in an arcuate delivery port ($5. One of the ports 65 is shown in Fig. 19. It will be noticed that both outlet ports are arcuate in shape and have a radial width equal to the diameter 13 of piston members 29b. Consequently the two ports overlie or coincide with a portion of the piston annulus as defined above.

One of the two cylinder blocks 15 receives an end of drive shaft 37 by which the cylinder blocli is rotated. This rotational movement is imparted through piston members 2917 to the other cylinder block. Both cylinder blocks are centrally recessed as at 44 and 33, one of them being counterbored at 3th: if necessary to give the central recesses the same diameter at the outer fades of the cylinder blocks.

In this form of pump, pressure relief passages are provided which function in the same manner as in the pumps described above, but the pressure relief passages take a slightly difierent form. Relief passage 43b in end plate 11b extends between central recess 44 and inlet passage 69. Consequently the fluid pressure within recess 44 is equal to the inlet pressure and any excess pressure is relieved through passage 43b. Any excess pressure building up in central recess 38, 33a is relieved through passage 45:; which is a longitudinally extending groove in the periphery of the bore through which drive shaft 37 passes. Alternatively, this bore in the end plate may be sulticiently oversized with respect to drive shaft 37 that pressure is relieved from recess 38 in the same manner. Relief passage 45]; extends from counterbore 33a into the chamber containing bearings 4d; and this bearing chamber is maintained at inlet pressure by virtue of relief passage 66 which places the bearing chain her in communication with fluid inlet passage til, as shown in Fig. 18. Passage 66 also provides means for introducing lubricant into bearings 40 when the lubricating fluid is being pumped.

It will be appreciated that with the construction of the pressure relief passage as shown in Fig. 18, that the pump is not reversible with respect to the direction of fluid flow through the pump unless the ehdplates are rotated if only shaft 37 were reversed in order tohave fluid enter the pump at 27!) and leave at 25b, the relatively, high outlet pressure would be established in. passages 60 61. This high pressure wduld pass through relief pas sages 43b and 66, building up the film pressure at the outer ends of the cylinder blocks and causing them to lose their seat on the end plate. As a result, the pump would fail to work properly. However, as mentioned above, end plates 11b and 1211 can be rotated through 180 if such change in direction of fluid flow is desired. Alternatively, duplicate passages for pressure relief communicating with outlet passages 62 and 63 may be provided and check valves inserted in the passages in the same general manner as disclosed in Figs. 13 and 14. The addition of these check valves would prevent the undesired reverse flow of. high pressure fluid and permit only the relief of pressure in the proper direction, as has already been explained.

Except for the changes specifically discussed, the construction of the pump shown in Figs. 17 and l9 is the same as that of the form first described.

In the event that it is not desired to permit fluid being pumped to have entry into the chamber containing shaft bearings 46), passages 45!) and 66 may be modified to permit pressure relief directly from recess 33 into inlet passage 61. This is in accord with the manner in which excess pressure in recess 34 is relieved through passage 4311. For example, relief passage 45b would extend only part way through end plate 12b and would then join passage 66 which would be relocated so that neither of the two passages communicates with the bearing chamber.

Having described a preferred form of my invention and certain modifications thereof, it will be apparent that various other changes in the construction and arrangement of the pump may be made by persons skilled in the art without departing from the spirit and scope of my invention. Consequently, 1 wish it understood that the foregoing is considered as illustrative of, rather than restrictive upon, the appended claims.

I claim:

1. A pump comprising a housing having a fluid inlet and a fluid outlet, rotary fluid displacement means of the angularly related piston and variable volume cylinder type including, a rotatable cylinder block having a sealing surface on one end intersected by a plurality of cylinder bores, means within the housing supporting the block for limited axial shifting in response to forces on the block, porting means providing a porting surface adjoining said sealing surface controlling the flow of fluid from said inlet to said bores and from said bores to said outlet, a pressure receiving surface on the other end of said cylinder block forming together with means including a portion of said housing a pressure space, means to fill said space with fluid at increased pressure generated by the pump thereby loading the cylinder block axially against said porting surface with thrust forces proportionate to pressures generated by the pump, recess means formed between said sealing surface and said port ing surface, and a passage extending from said recess means to a zone in said pump at low pressure for venting said recess means to a lower pressure than the operative pressures in said pressure space to maintain the forces acting on said cylinder block in the direction of said porting means proportionately greater than the forces acting on said cylinder block tending to separate said sealing and porting surfaces, whereby said sealing and porting surfaces are maintained in sealing relation.

2. A pump as defined in claim 1, wherein the means to fill said pressure space comprises running clearances between the moving parts of the pump.

3. A pump as defined in claim 1, wherein the means to fill said pressure space comprises a separate fluid passage formed in said housing.

4. pump as defined in claim 1 that also includes a shaft journalled in said housing and extending through said porting means and said recess means in driving connection with said cylinder block, the walls of said recess 16 means and said shaft defining an annular space therebetween circumscribing said shaft at the sealing and porting surfaces and constituting the recess means vented to said zone of lower pressure.

5. In a rotary fluid pump, the combination comprising: a housing having a fluid inlet and a fluid outlet and an internal cavity; a pair of cylinder blocks in the housing each rotatable about one of two angularly disposed intersecting axes and each block having on one end a pressure-receiving surface forming together with one another and with means including a portion of said housing a pressure space within the cavity, each block being supported for limited axial shifting within the cavity and each block having on the other end a sealing surface and each having a plurality of axially extending cylinder bores spaced apart in an annular row concentric with the block axis and intersecting said sealing surface; end plate means providing a porting surface adjoining each one of the sealing surfaces, at least one pair of fluid ports formed in said end plate means, each port being radially outward from a corresponding block axis and in register with an associated row of cylinder bores to control fluid flow from said inlet to said bores and thence to said outlet; a plurality of angular pistons slidably mounted in said cylinder bores; means to fill the pressure space with working fluid under a pressure generated by the pump for urging both the cylinder blocks axially outwardly toward the adjoining porting surfaces with thrust forces proportionate to pressures generated by the pump; a recess formed between each said sealing surface and the adjoining porting surface at a position radially inwardly of the cylinder bores; and passage means extending from said recesses to a Zone in said pump of low pressure for venting said recesses to lower pressure than the operative pressures in said pressure space to maintain the forces acting on said cylinder blocks in the direction of said porting means proportionately greater than the forces acting on said cylinder blocks tending to separate said sealing and porting surfaces, whereby said sealing and porting surfaces are maintained in sealing relation.

6. A pump as defined in claim 5, and a drive shaft journalled in one of said end plate means and having a portion extending through a corresponding one of said recesses and having a driving connection with a corresponding one of said cylinder blocks, the walls of said one recess and said shaft having an annular space therebetween circumscribing said shaft at the sealing and porting surfaces and constituting said vented recess at the drive end of the pump.

7. A pump as defined in claim 5 in which the piston members are hollow and the end plate means provide duplicate inlet ports in the two porting surfaces and an outlet port in one porting surface, the other porting surface having a recess corresponding in size and position to the outlet port.

8. A pump comprising a housing having an inlet and an outlet, rotary fluid displacement means of the angularly related piston and variable volume cylinder type including, a rotatable axially movable cylinder block having a first radially extending wall means including a sealing surface on one end intersected by a plurality of cylinder bores, porting means providing a second wall means adjoining said first wall means and said porting means including a porting surface adjoining said sealing surface for controlling the fiow of fluid from said inlet to said bores and to said outlet, said cylinder block having a pressure-receiving surface forming together with means including a portion of said housing a pressure space, means to fill said pressure space with fluid at increased pressure generated by the pump thereby pressure-loading said pressure-receiving surface and thrusting the cylinder block axially toward said porting means with forces pro-.

portionate to the pressures generated by the pump, recess means formed between said first and second wall means,

and a passage extending from said recess means to a zone in said pump at lower pressure than the operative pressures in said pressure space for venting said recess means, whereby the forces acting on said cylinder block in the direction of said porting means will be proportionately greater than the forces acting on said cylinder block tending to separate said sealing and porting surfaces, thereby to maintain said sealing and porting surfaces in sealing relation.

References Cited in the file of this patent UNITED STATES PATENTS 1,048,468 Brun Dec. 24, 1912 18 Nordberg Sept. 9, 1919 Rayfield July 12, 1932 Koplar Jan. 1, 1935 Thomas Apr. 9, 1935 Rayfield Mar. 16, 1937 Edmundson et al. May 25, 1937 Vickers May 26, 1942 Jones Mar. 2, 1943 Zimmermann July 18, 1944 MacNeil Dec. 5, 1944 FOREIGN PATENTS Italy 1931 Germany 1932 

